Part load operation of liquid air power and storage system

ABSTRACT

Apparatus, systems, and methods store energy by liquefying a gas such as air, for example, and then recover the energy by regasifying the liquid and combusting or otherwise reacting the gas with a fuel to drive a heat engine. The process of liquefying the gas may be powered with electric power from the grid, for example, and the heat engine may be used to generate electricity. Hence, in effect these apparatus, systems, and methods may provide for storing electric power from the grid and then subsequently delivering it back to the grid.

CROSS REFERENCE TO RELATED APPLICATIONS

This application is a continuation of International Patent Application No. PCT/US16/32342 titled “Part Load Operation of Liquid Air Power and Storage System” and filed May 13, 2016. PCT/US16/32342 claims benefit of priority to U.S. Provisional Patent Application No. 62/169,172 titled “Part Load Operation of Liquid Air Power and Storage System” and filed Jun. 1, 2015. Both of these applications are incorporated herein by reference in their entirety.

This application is related to U.S. patent application Ser. No. 14/546,406 titled “Liquid Air Power and Storage” and filed Nov. 18, 2014, which is incorporated herein by reference in its entirety.

FIELD OF THE INVENTION

The invention relates generally to storing energy by liquefying air or another gas or gaseous mixture, and subsequently recovering stored energy upon regasifying the liquid. More particularly the invention relates to regasifying the liquid, mixing the resulting gas with a fuel, and combusting or otherwise chemically reacting the gas-fuel mixture to drive a heat engine such as a turbine, for example, in a Liquid Air Power and Storage (LAPS) System, and to start-up, shut-down, normal full-power operation, and part load power generation of such a system.

BACKGROUND

The electric power system comprises generation, transmission, and distribution assets, which deliver power to loads. With the introduction of renewable resources, the electric power system faces a number of constraints which favor the addition of storage assets.

The principal constraint on an interconnected grid is the need to maintain the frequency and voltage by balancing variations in generation and demand load. Failure to maintain the voltage or frequency within specifications causes protective relays to trip in order to protect generators, transmission and distribution assets from damage. Because of the interconnected dynamic electrical grid, under frequency or undervoltage trips can cause a cascade of other trips, potentially leading to widespread blackouts.

Traditionally, electric utilities or the independent system operators managing electrical grids maintain a power generation reserve margin that can respond to changes in load or the loss of a generating unit or transmission line serving the load. These reserves are managed and scheduled using various planning methods, including day-ahead forecasts, dispatch queues that may be ordered based on generation cost, and generation ramp-rates, transmission constraints, outages, etc. The spinning generation units, that is, those that are operating, then respond to generation load control signals.

Many renewable resources are intermittent in nature, including wind farms, central station solar thermal or solar photovoltaic (PV) plants, and distributed photovoltaic systems including those on rooftops. These can produce power only when the resource is available, during daylight for solar, and when the wind is blowing for wind, leading to seasonal and diurnal production variations, as well as short-term fluctuations due to calms, gusts, and clouds. Gusts that exceed wind turbine ratings may cause them to trip with a sudden loss of full generation capacity. Deployment of these renewable systems as both central and distributed generating resources results in fluctuations in both the generation of power to be transmitted and the demand for power, since the distributed PV offsets load.

Base load is usually provided by large central station nuclear, hydroelectric or thermal power plants, including coal-fired steam plants (Rankine cycle) or gas-fired Combined Cycle Combustion Turbine plants (open Brayton air cycle with closed Rankine steam bottoming cycle). Base-load units often have operating constraints on their ramp rates (Megawatts per minute) and Turn-Down (minimum Megawatts), and startup from cold steel to rated load requires several hours to several days depending on the type and size of generating asset. Accordingly, a different class of load following power plants is also deployed in the electric power system, to complement the base load units. Generally, these load following units are less efficient in converting thermal energy to electrical energy.

This conversion efficiency is often expressed as a Heat Rate with units of thermal energy needed to produce a kilowatt-hour (kw-hr) of electricity [British Thermal Unit (BTU) per kw-hr in the U.S., kiloJoules (kJ) per kw-hr elsewhere]. The thermal equivalent of work is 3413 BTU/kw-hr or 3600 kJ/kw-hr, which represents 100% efficiency. Modern combined cycle power plants at full load rated conditions may have heat rates as low, for example, as 6000 kJ/kw-hr. Modern gas turbine peaking plants (e.g., General Electric LM6000-PC SPRINT) can achieve a full load rated condition heat rate of 9667 kJ/kw-hr (HHV). It is important to note that gas turbine heat rates increase rapidly away from rating conditions, and at part load in hot conditions the actual heat rate may be twice the rated Heat Rate.

It is of course desired to deliver electricity to customers at the lowest possible cost. This cost includes the amortization and profit on invested capital, the operating and maintenance (O&M) expense, and the cost of fuel. The capital amortization (and return on capital, in the case of regulated generators) is applied to the capacity factor (fraction of rated generation) to arrive at the price ($ per Megawatt-hour) associated with the fixed capital expense. The Heat Rate multiplied by the fuel cost determines the contribution of the variable fuel consumption to the electricity price. The O&M expenses generally have some combination of fixed and variable expenses, but are insignificant compared to capital and fuel for central station plants. Generating units have different mixes of fixed and variable expenses, but presumably were believed to be economic at the time they were ordered.

In order to deliver low cost electricity to a customer, it is necessary to operate the capital intensive units at high capacity, subject to fuel cost, in order to spread the capital cost across many kw-hr. Contrariwise, it is necessary to minimize the operation of units with high marginal operating cost (due to high Heat Rate, Fuel Cost or O&M). This was indeed the planning assumption for procurement of the existing fleet of generators.

The Renewable resources gather ‘free’ fuel, so their cost of generation is dominated by the amortization of the capital needed to gather and convert this energy into electricity. In order to profitably build and operate a Renewable power plant, it should have as high a capacity factor as may be practically realized. Similarly, the fuel-efficient base load generation should operate at high capacity factor, both to amortize the capital expense, and because its operating characteristics induce higher fuel or O&M costs (per unit of generation) when operated intermittently or at part load.

The increasing penetration of renewable generation with variable generation characteristics is challenging the traditional dispatch order and cost structure of the electric generation system. In practice, utility scale solar power plants without storage are limited to Capacity Factors of about 25%, and wind farms seldom exceed 50%. This capacity may not coincide with demand, and may be suddenly unavailable if the sun or wind resource is reduced by local weather. For example, if wind resources are available at periods of low demand, base load units must either ramp down or shut-down or the wind resource must be curtailed. If the wind is not curtailed, then less efficient peaking units may be needed to provide ramp flexibility that the large base-load units cannot provide in case of gusts or calms. Likewise the widespread deployment of solar power generation is depressing the need for generation during daylight hours, but large ramp rates as the sun rises and sets can currently only be met by gas fired peaking plants. Ironically, this will result in displacement of low-cost, high efficiency base-load units in favor of high cost, low-efficiency peaking units, with a concomitant increase in greenhouse gas releases.

For environmental, energy security, cost certainty and other reasons, renewable energy sources are preferred over conventional sources. Demand Response techniques, which attempt to reduce the instantaneous load demand to achieve balance between generation and consumption, are analogous to a peaking generation unit. Another approach is deployment of (e.g., large scale) energy storage systems which would mediate the mismatch between generation and consumption.

Storage systems are alternately charged to store energy (e.g., using electric power), and discharged to return the energy as power to the electric grid. The technical characteristics of energy storage systems include:

-   -   the Capacity, or quantity of energy that can be stored and         returned, measured in MW-hours;     -   the Round Trip Efficiency (RTE), or fraction of the energy         delivered to the storage system that is returned to the grid;     -   the Power rating, or rate in MW at which the system can be         charged or discharged (Power is often symmetric, though this is         not necessary, or even desirable);     -   the Lifetime, which is typically the number of Charge/Discharge         cycles.

Pumped Storage Hydroelectricity (PSH) employs a reversible pump-turbine with two water reservoirs at different elevations. When excess energy is available, it is used to pump water from a lower to an upper reservoir, converting the electricity into potential energy. When electricity is needed, the water flows back to the lower reservoir through a hydro-turbine-generator to convert the potential energy into electricity. Pumped hydro storage is suitable for grid scale storage and has been used for many decades in electrical grids around the world. PSH has a Round Trip Efficiency (RIE) of 70% to 80% and can be deployed at Gigawatt scale with many days of potential storage. In addition to high RIE, PSH does not generate greenhouse gases during operation. Deployment of PSH requires availability of suitable locations for the construction of dams and reservoirs, and evaporative water loss may be an issue in some locations.

Compressed Air Energy Storage (CAES) stores pressurized air that is subsequently expanded in an engine. Commercially deployed CAES stores the air in large underground caverns such as naturally occurring or solution-mined salt domes, where the weight of overburden is sufficient to contain the high pressures. The RTE for CAES may be relatively low. The 110MW McIntosh CAES plant in the US state of Alabama reportedly has a RTE of only 27%, for example. Several near-isothermal CAES technologies are also under development with reported RTE of 50% or greater, using pressure vessels for storage.

Many energy storage technologies are being developed and deployed for end-use loads or distribution level capacities, at power levels from a few kilowatts to several megawatts. These approaches typically employ batteries with a variety of chemistries and physical arrangements.

There is a need for high efficiency energy storage that is not dependent on geological formations, and which can be deployed at scales of tens to hundreds of megawatts to complement the existing generation and transmission assets.

SUMMARY

Apparatus, systems, and methods described in this specification store energy by liquefying a gas such as air, for example, and then recover the energy by regasifying the liquid and combusting or otherwise reacting the gas with a fuel to drive a heat engine. The process of liquefying the gas may be powered with electric power from the grid, for example, and the heat engine may be used to generate electricity. Hence, in effect these apparatus, systems, and methods may provide for storing electric power from the grid and then subsequently delivering it back to the grid.

One aspect of the invention is a method of operating a liquid air power and storage system comprising a combustor and a turbine. The method comprises regasifying liquid air or liquid air components to produce gaseous air or gaseous air components, combusting at least a portion of the gaseous air or gaseous air components in a combustor with a fuel (e.g., a gaseous fuel) to form a gaseous primary working fluid at an elevated temperature, expanding the primary working fluid through the turbine, and adjusting a mass flow rate of the gaseous air or gaseous air components into the combustor proportionally to a changing mechanical load on the turbine. The LAPS system may include, for example, a generator driven by the turbine to generate electricity and thereby providing the changing mechanical load on the turbine. Alternatively, the changing mechanical load on the turbine may be provided by, for example, a pump drive, a compressor drive, or a drive providing motive power for propulsion. This method may be used, for example, to efficiently operate the liquid air power and storage system at a power output less than a maximum output power for the system.

The method may comprise regasifying the liquid air or liquid air components to produce gaseous air or gaseous air components using heat from an exhaust gas stream from the turbine, but this is not required. Heat derived from other sources (i.e., not derived from the turbine exhaust stream) may be used in addition to or instead of heat from the turbine exhaust gas stream to regasify the liquid air or liquid air components.

The method may comprise adjusting a mass flow rate of the fuel into the combustor to maintain a desired temperature of the primary working fluid at an inlet or an outlet to the turbine.

If the LAPS system comprises a generator driven by the turbine, the method may comprise operating the generator as a synchronous condenser.

The method may comprise operating the liquid air power and storage system with the mass flow rate of the gaseous air or gaseous air components into the combustor less than that required to operate the turbine at nameplate capacity, a mass flow rate of the fuel into the combustor less than that required to operate the turbine at nameplate capacity, and a temperature of the primary working fluid at an inlet to the turbine measured in degrees Kelvin at least 60% of that required to operate the turbine at nameplate capacity. In such cases the temperature of the primary working fluid at the inlet to the turbine measured in degrees Kelvin may be, for example, at least 65%, at least 70%, at least 75%, at least 80%, at least 85%, at least 90%, at least 95%, or at least 100% of that required to operate the turbine at nameplate capacity. The mass flow rate of the gaseous air or gaseous air components into the combustor may be, for example, less than 95%, less than 90%, less than 85%, less than 80%, less than 75%, less than 70%, less than 65%, less than 60%, less than 55%, less than 50%, less than 45%, less than 40%, less than 35%, less than 30%, less than 25%, less than 20%, less than 15%, less than 10%, or less than 5% of that required to operate the turbine at nameplate capacity.

The method may comprise operating the liquid air power and storage system with the mass flow rate of the gaseous air or gaseous air components into the combustor greater than that required to operate the turbine at nameplate capacity.

The method may also comprise producing the liquid air or liquid air components in an electrically powered liquefaction process and storing the liquid air or liquid air components for later regasification and use in combusting the gaseous fuel. As elaborated on below, heat from the turbine exhaust gas stream used in regasifying the liquid air or liquid air components may be transferred from the turbine exhaust gas stream to the liquid air or liquid air components via a secondary working fluid, but this is not required.

The fuel and gaseous air or gaseous air components may be combusted using a flame, or catalytically combusted. A gaseous fuel may be or comprise, for example, natural gas, hydrogen gas, or a mixture of natural gas and hydrogen gas. A liquid fuel may be a petroleum distillate, for example. Different ones of the gaseous air components may be separately provided to a combustor in which they are combusted with the fuel, for example to support oxy-combustion followed by nitrogen dilution. The gaseous primary working fluid may be provided to an inlet of the turbine at, for example, a temperature greater than or equal to about 1600 K and a pressure greater than or equal to about 30 bar.

The method may comprise heating water to produce steam using heat from the exhaust gas stream from the turbine, and mixing some or all of the steam with the gaseous air or gaseous air components and fuel during combustion of the fuel. In addition, or alternatively, the method may comprise heating water to produce steam using heat from the exhaust gas stream from the turbine, and using some or all of the steam to cool the turbine. In addition, or alternatively, the method may comprise cooling the turbine with a portion of the gaseous air or gaseous air components provided to the turbine separately from the primary working fluid. In the latter case, the liquid air or liquid air components may be regasified in a regasifier using heat from an exhaust gas stream from the turbine, and the gaseous air or gaseous air components provided to the turbine to cool the turbine may be drawn, for example, from an intermediate location in the gasifier at a temperature lower than a maximum temperature of the gaseous air or gaseous air components at an outlet of the gasifier. The method may comprise decreasing the mass flow rate of the gaseous air or gaseous air components into the combustor in combination with increasing a flow rate of the portion of gaseous air or gaseous air components provided to the turbine separately from the primary working fluid to cool the turbine.

The method may comprise heating a secondary working fluid with heat from the exhaust gas stream from the turbine to convert the secondary working fluid from a liquid phase to a gas phase, and using the heated working fluid in a bottoming Rankine power cycle to generate additional mechanical power (e.g., to drive a generator to generate electricity). In such cases, the method may comprise proportionately varying a flow rate of the secondary working fluid and the mass flow rate of the gaseous air or gaseous air components into the combustor. The method may comprise operating the bottoming Rankine power cycle in sliding pressure mode. The secondary working fluid may be an organic working fluid, for example, and the method may comprise regasifying the liquid air or liquid air components using heat transferred from the organic secondary working fluid to the liquid air or liquid air components, thereby condensing the organic secondary working fluid from a gas phase to a liquid phase. Alternatively, the secondary working fluid may be or comprise water.

The method may comprise freezing carbon dioxide out of the exhaust gas stream from the turbine by transferring heat from the turbine exhaust gas stream to the liquid air or liquid air components to cool the turbine exhaust gas stream.

The method may comprise condensing water out of an exhaust gas stream from the turbine thereby increasing the turbine power output.

These and other embodiments, features and advantages of the present invention will become more apparent to those skilled in the art when taken with reference to the following more detailed description of the invention in conjunction with the accompanying drawings that are first briefly described.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 shows a block diagram of an example liquid air power and storage system utilizing hybrid regasification to recover the stored energy for electricity production.

FIGS. 2-4, 6, 11, 13, and 16 show block diagrams of additional example liquid air power and storage systems utilizing hybrid regasification to recover the stored energy for electricity production.

FIG. 5 shows a plot illustrating the effect of steam injection on the efficiency of electric power production by the example liquid air power and storage system of FIG. 4.

FIG. 7 shows a plot illustrating an example charge/discharge cycle for a liquid air power and storage system as disclosed herein, in which charging and discharging occur at different times.

FIG. 8 shows a plot illustrating an example charge/discharge cycle for a liquid air power and storage system as disclosed herein, in which the charging system operates continuously.

FIG. 9 shows a block diagram of a conventional combustion turbine-generator system.

FIG. 10 shows a plot (percent of full load parameter versus percent of full load power) illustrating part load performance for a conventional combustion turbine-generator system as illustrated in FIG. 9, for example.

FIG. 12 shows a plot (percent of full load parameter versus percent of full load power) illustrating part load performance for an example LAPS system as illustrated in FIG. 11, for example.

FIG. 14 shows a plot (percent of full load parameter versus percent of full load power) illustrating part load performance for an example LAPS system as illustrated in FIG. 13, for example

FIG. 15 shows a plot of HHV Heat Rate versus power for an example combustion turbine generator, an example simple LAPS system, and an example LAPS system using steam injection.

DETAILED DESCRIPTION

The following detailed description should be read with reference to the drawings, in which identical reference numbers refer to like elements throughout the different figures. The drawings, which are not necessarily to scale, depict selective embodiments and are not intended to limit the scope of the invention. The detailed description illustrates by way of example, not by way of limitation, the principles of the invention. This description will clearly enable one skilled in the art to make and use the invention, and describes several embodiments, adaptations, variations, alternatives and uses of the invention, including what is presently believed to be the best mode of carrying out the invention. As used in this specification and the appended claims, the singular forms “a,” “an,” and “the” include plural referents unless the context clearly indicates otherwise.

This specification discloses apparatus, systems, and methods by which energy may be stored by liquefying a gas such as air, for example, and then recovered by regasifying the liquid and combusting the gas with a fuel to drive a heat engine. The combination of regasifying the liquid and combusting it with a fuel is referred to herein as “hybrid regasification”. The process of liquefying the gas may be powered with electric power from the grid, for example, and the heat engine may be used to generate electricity. Hence, in effect these apparatus, systems, and methods may provide for storing electric power from the grid and then subsequently delivering it back to the grid. The electricity for liquefying the gas may be provided, for example, from base load power generation or from renewable resources that would otherwise be curtailed, and hence may be effectively low cost. Consuming the fuel in this manner may increase the efficiency with which the fuel is converted to electricity, compared to conventional power generation.

Referring now to example liquid air energy storage system 5 illustrated by the block diagram of FIG. 1, liquid air (or another liquefied gas) from liquid air source 10 is converted to a gaseous state in regasifier 20 and provided to heat engine 40. Fuel from fuel source 30 is also provided to heat engine 40, where it is combusted or otherwise chemically reacted with the gas from regasifier 20 to generate heat. Heat engine 40 converts the heat to work that is used to drive electric generator 50 to generate electricity.

Liquid air source 10 may comprise, for example, apparatus for liquefying air or another gas, and storage tanks for storing the resulting liquid. Alternatively, or in addition, liquid air or another liquefied gas may be produced off-site and delivered by tanker truck, tanker rail car, or any other suitable transport to storage tanks in liquid air source 10. Hence liquid air source 10 need not include liquefaction apparatus. Although in examples described below this specification refers to the liquefied gas used in the energy storage systems it describes as liquid “air” (i.e., the mixture of nitrogen, oxygen, and other gases occurring naturally in the atmosphere near the surface of the earth) or liquefied air components, any suitable gas or mixture of gases that may be chemically reacted with a fuel to produce heat may be used in addition or instead of liquid air. The liquefied gas may be stored at approximately atmospheric pressure, for example.

Fuel source 30 may comprise any suitable combination of fuel storage, fuel conditioning, and fuel transport apparatus. The fuel used in the energy storage systems described herein may be or comprise, for example, any suitable gaseous, liquid, or solid material that may be chemically reacted with gas from regasifier 20 to generate heat. Suitable fuels include but are not limited to, for example, natural gas, hydrogen, petroleum, petroleum products, petroleum coke, coal, pulverized coal, wood, or other naturally occurring or man-made carbonaceous materials. The fuel may be a fossil fuel, for example. Solid or liquid materials may, for example, be gasified or converted to another gaseous fuel (for example, syngas) prior to delivery to heat engine 40. Alternatively, or in addition, solid or liquid fuels may be delivered as solids or liquids to heat engine 40.

Heat engine 40 may combust or catalytically combust the gas-fuel mixture, or use any other suitable chemical reaction of the gas-fuel mixture to produce heat. Heat engine 40 may be or comprise a combustion turbine, a steam turbine, an organic working fluid gas turbine, a reciprocating engine, or any other suitable heat engine. Heat engine 40 may operate on any suitable thermodynamic cycle. As further described below, liquid air energy storage systems as described herein may optionally include more than one heat engine, in which case the heat engines may be arranged with the exhaust of one heat engine used as the heat source for another heat engine in a combined cycle configuration.

Generator 50 may be or comprise any suitable electric generator.

FIG. 2 shows a block diagram of an example liquid air energy storage system 99 that includes subsystems corresponding individually or in combination to the subsystems of energy storage system 5 of FIG. 1, as well as additional subsystems. Liquid air energy storage system 99 is described below as combusting a gaseous fuel (e.g., natural gas) in the presence of regasified liquid air or regasified liquid air components to drive a combustion turbine generator. As the preceding portion of the specification indicates, however, the invention is not limited to this particular combination of liquefied gas, fuel, and heat engine. Rather, any suitable combination of liquefied gas, fuel, and heat engine may be used in variations of liquid air energy storage system 99.

Referring now to FIG. 2, during charging of liquid air energy storage system 99 gaseous air is liquefied in air liquefaction subsystem 100, which consumes electricity to liquefy air using any suitable commercially available air liquefier. Air liquefiers and related technologies suitable for use in this subsystem may include those available, for example, from Air Liquide S. A., Praxair Technology, Inc., and Linde AG. Heat evolved during the liquefaction process as a result of the phase change from gas to liquid may optionally be stored in an optional thermal energy storage subsystem (not shown) as sensible heat or as latent heat in suitable phase change materials, for subsequent use in regasifying the liquid air in liquid air regasification subsystem 400 described in more detail below. Alternatively or in addition, some, most, or all of the heat removed during the liquefaction process may be rejected to the environment.

Air liquefaction subsystem 100 may be designed to separate the air into its various components. Some of the components, such as Argon, may be separated out and sold for use by third parties. As another example, carbon dioxide may be separated out for sale and/or use in enhanced oil recovery, or otherwise sequestered as part of a greenhouse gas reduction program. Fractionation of air may also be used for example for the principal components, nitrogen and oxygen, to permit oxy-combustion (in the heat engine) to reduce nitrogen oxide formation, facilitate sequestration of carbon dioxide, and/or to permit commercial sale of these air components if they are not needed in their naturally occurring amounts or ratio.

The liquefied air or liquefied air components produced by air liquefaction subsystem 100 may be transferred to and stored in liquid air storage subsystem 200, which typically comprises one or more conventional insulated storage tanks maintained near atmospheric pressure. The cryogenic liquid may be stored as a mixture or as separate components in different tanks. The quantity of liquefied air in the tanks increases as the energy storage system is charged, and decreases as the energy storage system is discharged. Although there are small evaporative losses as a result of heat transfer from the environment, the cryogenic liquid can be stored for extended periods, facilitating energy storage during weekends, at night, or during other periods when, for example, relatively less expensive electricity is available in anticipation of generation during periods of higher value. The liquefied air may be stored at a temperature of about 78 K, for example.

During discharge of liquid air energy storage system 99, one or more liquid air pumps 300 are controlled to pump liquid air from liquid air storage subsystem 200 in response to the demand for energy. If the liquid air is stored as a mixture of components to be transferred together, a single pump 300 (as shown) may be sufficient. If instead different liquid air components are to be separately transferred, to facilitate oxy-combustion for example, two or more pumps 300 may be arranged in parallel to do so. The pump or pumps may operate at variable speeds to provide a metering function and/or to maintain a desired or necessary pressure at an inlet to combustion turbine generator subsystem 700. Alternatively, the pump or pumps may operate at constant speeds, and control valves (not shown) may be used to meter the liquid air and/or to maintain the pressure at the inlet to the turbine generator.

Pump or pumps 300 may pump the liquefied air, which is typically at about atmospheric pressure during storage, to a higher inlet pressure required by the heat engine. For example, pump 300 may pump the liquefied air to a pressure of about 35 bar.

Liquid air pumped by pump 300 from liquid air storage subsystem 200 passes through liquid air regasification subsystem 400, which evaporates the cryogenic liquid by heat transfer from, for example, working fluids, combustion heaters, and/or thermal energy storage to provide high pressure gaseous air. (Regasification subsystem 400 is also referred to in this specification as “regasifier 400”). The heat used in this regasification step may be transferred (e.g., using one or more working fluids) from the exhaust gas stream from the turbine, but this is not required. Heat derived from other sources (i.e., not derived from the turbine exhaust stream) may be used in addition to or instead of heat from the turbine exhaust gas stream to regasify the liquid air or liquid air components. For example, in addition to working fluids, combustion heaters, and thermal energy storage as noted above, other heat sources for the regasification step may include ambient air and/or process heat from outside the LAPS system. In the illustrated example, during normal operation most of the heat for this process is supplied by heat exchange from the sensible and latent heat of the working fluid used in an organic working fluid Rankine bottoming cycle described in more detail below. As an alternative (not shown in FIG. 2), heat for liquid air regasification subsystem 400 may be discharged from heat recovery steam generator subsystem 800, which is also described in more detail below. High pressure liquid air may enter regasification system 400 at a pressure of about 35 bar and temperature of about 80 K, for example. High pressure gaseous air may leave regasification system 400 at a pressure of about 35 bar and a temperature of about 380 K, for example.

Auxiliary air heater 410, which may be or include one or more combustion heaters for example, and/or may include heat exchangers using heat available from other processes, may provide auxiliary heat to the liquid or gaseous air at system start-up or during other transients, such as load changes, for which heat transfer to regasification subsystem 400 from the bottoming cycle or from the heat recovery steam generator 800 may be insufficient. Auxiliary air heater 410 may also be used at other times as well, to stabilize and/or control the temperature of the gaseous air provided to the heat engine and to increase its temperature. High pressure gaseous air may leave auxiliary air heater 410 at a pressure of about 35 bar and a temperature of about 400 K to about 500 K, for example. Auxiliary air heater 410 may be bypassed or not operated so that high pressure gaseous air is provided to the heat engine at approximately the same conditions as produced in the regasifier 400.

After passing through auxiliary heater 410, the high pressure gaseous air enters combustion subsystem 600, where it is mixed and combusted with fuel from fuel supply and conditioning subsystem 500. (Combustion subsystem 600 is also referred to in this specification as “combustor 600” or “burner 600”).

Although the illustrated examples show auxiliary heater 410 located between gasifier 400 and combustor 600, auxiliary heater 410 may alternatively be located between cryopump 300 and gasifier 400.

Fuel supply and conditioning subsystem 500 conditions and meters the fuel in order to regulate the temperature at the inlet to the combustion turbine generator subsystem 700, which is described in more detail below. In the illustrated example the fuel is a gas, such as natural gas or hydrogen, for example. Alternatively, the fuel may be a liquid, for example a distillate fuel such as an oil or alcohol, or a solid such as coal or wood, for example. As suitable, customary compressors, pumps, or conveyors are used to condition and meter the fuel. In the illustrated example, the gaseous fuel may be supplied to fuel system 500 from tanks or by pipeline at, for example, atmospheric pressure and ambient temperature. Fuel system 500 compresses the gas to raise its pressure and temperature to, for example, about 31 bar and about 667 K, and supplies it to combustion system 600. In other embodiments the gaseous fuel may be supplied at higher pressures, and require little or no pressure increase.

Combustion subsystem 600 mixes and combusts the air and fuel supplied from regasification subsystem 400 and fuel system 500 to create a high temperature and high pressure primary working fluid for expansion through combustion turbine generator subsystem 700. In the illustrated example, combustion subsystem 600 also mixes in superheated steam, but this is not required. The primary working fluid comprises uncombusted regasified air components (principally nitrogen and oxygen), the products of combustion (principally water vapor and carbon dioxide), and may also include superheated steam. It is often desired to reduce the production of nitrogen oxides, which are a regulated pollutant and precursor to photochemical smog. In some variations, the air, fuel and superheated steam may be pre-mixed in various combinations to reduce the unwanted production of nitrogen oxides. In other variations, the fuel may be burned in a pure oxygen environment, followed by steam injection to reduce the temperature, followed by dilution with nitrogen. This also may reduce the unwanted production of nitrogen oxides. Sufficient reduction of nitrogen oxides may permit less expensive pollution control measures, such as operation without expensive selective catalytic reduction. In some variations, a catalytic combustor may be used instead of a flame. The heat of combustion increases the temperature of the mixture of air, steam and combustion products to, for example, about 1600 K, which is a typical inlet temperature for modern combustion turbines.

The hot gaseous primary working fluid created in combustion subsystem 600 is provided to the inlet of combustion turbine generator subsystem 700, which extracts energy from the working fluid as it expands through the turbine. The turbine may be similar in design and construction to modern combustion turbine power sections. The primary fluid may have, for example, a temperature of about 1600 K and a pressure of about 31 bar at the inlet to the turbine. Optionally, a stream of regasified air from regasification subsystem 400 may bypass combustion subsystem 600 and be provided to turbine generator subsystem 700 to cool the turbine components, including blades and vanes. This may allow operation at temperatures up to about 1750 K or higher, for example. Alternatively or additionally, the turbine blades may be optionally cooled with steam. Other temperature and pressure combinations may be used and will result in different flow rates throughout the overall system. Generally, the pressure and temperature may be selected to optimize the economic return, subject to constraints of materials and code classifications.

In the illustrated system, high turbine inlet pressure is available without the high compression work required in a conventional Brayton cycle combustion turbine generator (which includes a compressor section) because both the liquid air and the liquid water (vaporized in the heat recovery steam generation subsystem 800) are approximately incompressible fluids. Because performance is less sensitive to pressure drop, this also permits more flexible arrangement of the combustion subsystem inlet ducting and power take-off.

The turbine section of the turbine generator may, for example, be coupled to a synchronous generator, either directly or through a gearbox. In the synchronous connection, rotational speed and volumetric flow is constant, and because flow is choked, the inlet pressure is proportional to the power output. Accordingly the outlet pressure of the liquid air, liquid water and gaseous fuel supply may be regulated in proportion to the output power.

FIG. 2 shows combustion subsystem 600 and combustion turbine generator subsystem 700 separately. In some variations these subsystems may be physically separate, which may facilitate use of solid fuels. Alternatively, these subsystems may be physically integrated, in which case they may be viewed as components of a single subsystem.

The primary working fluid pressure and temperature is reduced after energy is extracted by combustion turbine subsystem 700, for example to about 800 K and about 1.5 bar. The turbine generator discharge pressure is the “back pressure” necessary for the exhaust mixture to flow through heat recovery steam generator subsystem 800, organic vapor generator 1010, and flue gas condensate collector 920 and its flue. It may be desirable to minimize the back pressure in order to extract more work from the primary working fluid. Condensing water vapor from the exhaust mixture induces a vacuum effect which tends to reduce the back pressure, thereby increasing the extracted work.

In the illustrated example, after exiting the combustion turbine the primary working fluid flows across a heat exchanger in heat recovery steam generator subsystem 800. Boiler feedwater is pumped through this heat exchanger to produce the superheated steam which is mixed with the fuel and air in combustion subsystem 600. The superheated steam may have a pressure of about 35 bar and a temperature of about 780 K, for example. The primary working fluid, which has now been cooled, for example, to about 578 K at about 1.25 bar, then flows across a heat exchanger in organic vapor generator 1010 to heat a second (organic) working fluid. On exiting this second heat exchanger the primary working fluid flows into flue gas condensate collector 920. At this stage the primary working fluid is cooled sufficiently to condense water vapor, which is separated and extracted, with the balance of the gas exhausted up a flue or chimney.

In an alternative variation, the flue gas may be further cooled to approximately 190 K, for example, using the liquid air stream in order to freeze carbon dioxide. A scraped surface heat exchanger would allow crystalline carbon dioxide to be collected for capture and storage. The carbon dioxide could then be sequestered, thereby permitting the system to be carbon dioxide neutral, despite burning fossil fuel. Alternatively, the solid-phase carbon dioxide could be stored and then used to pre-cool air in the liquefaction system, with the sublimated carbon dioxide being released to the atmosphere, captured as a carbonate, or compressed for sequestration as a vapor.

A selective catalytic reduction (SCR) system may be located at any suitable place in the exhaust stream from combustion turbine generator subsystem 700 in order to reduce the concentration of nitrogen oxides to permissible levels.

Water condensed from the primary working fluid in flue gas condensate collector 920 may be separated and collected in a tank. The condensate arises from superheated steam that was mixed with the fuel and air in combustion subsystem 600 and from water of combustion (a reaction product resulting from hydrogen in the fuel). This condensate may require treatment before being returned to heat recovery steam generator subsystem 800, for example to adjust the pH because of carbonic or nitrous/nitric acids that may dissolve in the condensate from carbon dioxide or nitrogen oxide combustion products.

The larger fraction of condensate is recycled to the heat recovery steam generator subsystem 800 by boiler feedwater pump 930, which increases the pressure of the boiler feedwater sufficiently for the resulting superheated steam to be mixed in combustion subsystem 600. The flow rate of feedwater may be controlled, for example, to maintain a specified superheated steam temperature, in order to maximize the economic return during the energy discharge mode of operation. This economic return may be a balance between fuel cost and electricity value, such that increased feedwater flow necessitates increased fuel flow (at constant combustion system outlet temperature) resulting in increased power output, versus constant fuel flow with higher feedwater flow and lower temperature.

Heat recovery steam generator subsystem 800 may have a once through configuration, which is capable of rapid startup and shutdown, high turn-down ratio and dry operation (without feedwater flow or steam generation) to provide the operating flexibility desired in an energy storage system that may be dispatched quickly and at varying loads. Alternative variations may instead employ drum type boilers with either natural or forced circulation. Boiler feedwater pump 930 may be controlled by a variable speed drive to deliver liquid water to the heat recovery steam generator, or a control valve may be used. In a once through configuration, feedwater is metered to control the temperature of superheated steam leaving the boiler, it typically being desirable to maximize the temperature in order to minimize fuel consumption. In a drum boiler configuration the feedwater is metered to maintain water inventory, i.e. the drum level, and a steam control valve is used to meter the steam flow rate in order to manage the steam drum pressure.

Liquid air energy storage system 99 includes an optional Rankine bottoming cycle that extracts heat from the primary working fluid to a suitable secondary working fluid such as, for example, an organic, water, ammonia, or refrigerant to generate additional electric power, while transferring heat to regasification subsystem 400. The secondary working fluid may be a binary mixture and/or an azeotrope, for example. In the illustrated example, regasification subsystem 400 condenses a low pressure gaseous organic secondary working fluid to produce a low pressure organic working fluid. Organic liquid pump 1030 then pumps the liquid to a higher pressure and through organic vapor generator 1010, which vaporizes the liquid to provide a high pressure gas that expands through a turbine in organic turbine generator 1000 to generate additional electric power. The low pressure organic vapor exhaust from organic turbine generator 1000 is then recycled through regasification subsystem 400. Suitable organic Rankine cycle turbine generators may include those available from Ormat Technologies, Inc. Suitable organic working fluids may include, for example, alkanes (e.g., propane) and hydrofluorcarbons (e.g., R134a), or other compounds such as ammonia or carbon dioxide.

A startup heater 1020, which may be or include one or more combustion heaters for example, and/or may include heat exchangers using heat available from other processes, may provide auxiliary heat to the liquid or gaseous organic fluid at system start-up or during other transients, such as load changes, for which heat transfer to regasification subsystem 400 from the bottoming cycle or from the heat recovery steam generator may be insufficient. Startup heater 1020 may also be used at other times as well, to stabilize and/or control the temperature of the secondary working fluid to the organic turbine generator 1010 and to increase its temperature. Startup heater 1020 may share a heat source with auxiliary air heater 410 described above. Optionally, the liquid air power and storage system may include only one or the other of startup heater 1020 and auxiliary heater 410, but not both, as their functions may be duplicative or partially duplicative. As described later, the system can be started and synchronized with a small fraction of the rated air flow, so the size of startup heater and the starting fuel requirements may be minimized. The system may also be started using compressed air through the startup heater or combustor, to bootstrap the system and provide heat for gasification of liquid air.

If propane is used as the secondary working fluid, it may be heated for example to approximately 575 K at about 100 bar, expanded through the turbine generator to generate electric power, and then condensed at about 1.5 bar in the regasification subsystem to transfer heat to the evaporating cryogenic air. High pressure gaseous propane may leave startup heater 1020 at a pressure of about 100 bar and a temperature of about 500 K to about 600 K, for example.

The use of an organic Rankine bottoming cycle employing regasification subsystem 400 as a heat sink may be particularly advantageous, because the organic working fluid may be selected to condense but not freeze during heat exchange with liquid air in regasification subsystem 400. In contrast, a Rankine bottoming cycle employing water as a working fluid and regasification system 400 as a heat sink might risk freezing the working fluid.

Variations of liquid air energy storage system 99 may employ a steam Rankine bottoming cycle in addition to or instead of the illustrated organic Rankine cycle, however. For example, a steam Rankine cycle may be inserted between heat recovery steam generator subsystem 800 and the organic Rankine cycle subsystem, with heat exchange between the steam and organic Rankine bottoming cycles providing a heat sink for the steam Rankine cycle and a heat source for the organic Rankine cycle.

Tables 1-5 below identify unit operations and provide process conditions for an example implementation of liquid air storage energy system 99 modeled using conventional commercially available power engineering software tools as if it were a continuous process. In the modeled implementation, for which the fuel is natural gas, an air flow of about 100 kilograms per second would generate approximately 120 MW of electric power while consuming about 168 MW of fuel (on a Higher Heating Value basis). If air was liquefied only during the discharge phase, approximately 88 MW of electric power would be consumed, resulting in an overall RTE of about 47%. Assuming six hours of discharge, there would be four hours of charging for each hour of discharge, so the liquefaction system would consume approximately 22 MW. The modeled process assumes perfect combustion without unburned hydrocarbons or production of nitrogen oxides or carbon monoxide. Also, in the modeled implementation the liquid air is not separated into its components.

Typical operation of such an energy storage system might demand, for example, 4 hours of discharge, requiring 14,400 seconds times 100 kg/s or 1440 metric tonnes of liquid air. This is at the small range of tonnage of commercially available air separation plants, for example. At a specific volume of 0.001147 cubic meters per kilogram, 1650 cubic meters of storage tanks would be required. This could be satisfied by two ten meter high by ten meter diameter atmospheric storage tanks. The tanks may typically be filled during periods of low electricity prices, for example on weekends and overnight. For redundancy and to take advantage of periods of low capacity or low electric prices, extra tanks may be installed.

The process conditions listed in Tables 1-5 are intended to be illustrative but not limiting. Other temperatures, pressures, flow rates, and working fluids may be utilized in order to optimize the cost, auxiliary power consumption, safety and handling, or convenience.

FIG. 3 shows an example liquid air power and storage system that includes additional optional components, compared to the system of FIG. 2, that provide alternative bottoming cycle arrangements. In the example system of FIG. 3 valves 1040, 1041, and 1042 may be configured to optionally route the bottoming cycle working fluid to bypass organic turbine generator 1000 and instead flow directly to regasification system 400. This bypass configuration may be useful, for example, during startup of the bottoming cycle or when the bottoming cycle turbine generator is out of service.

The example system of FIG. 3 also includes an optional recuperator 1050 positioned in the exhaust stream of organic vapor generator 1000 upstream from regasification system 400. Recuperator 1050 preheats the high pressure organic liquid exiting organic liquid pump 1030 with heat transferred from the low pressure organic vapor exhaust from organic turbine generator 1000, before the organic liquid enters organic vapor generator 1010.

In alternative variations organic turbine generator 1000 and recuperator 1050 are absent, and an organic working fluid circulates to transfer heat from the combustion turbine exhaust to regasification system 400 in a flow path similar or equivalent to the organic turbine generator bypass configuration of the system shown in FIG. 3.

Referring again to FIG. 3, the illustrated system also comprises optional condensate tank 940, water treatment system 950, and treated water tank 960. Water may flow from condensate collector 920 to condensate tank 940. Water treatment system 950 may process the condensate to remove oxides of carbon, nitrogen, sulfur, or other combustion products, and to adjust pH for storage in treated water tank 960. Boiler feedwater pump 930 draws water from treated water tank 960. Tanks 940 and 960 may each be sized to accommodate the total water condensed during discharge of the energy storage system. Water Treatment System 950 may optionally be operated exclusively or principally during the charging phase, to reduce energy demand during the more valuable discharge phase and to permit selection of a smaller capacity and hence cheaper water treatment system which could process condensate during the longer duration charging phase.

FIG. 4 shows an example liquid air power and storage system that lacks a bottoming cycle and uses the combustion turbine exhaust gas to directly heat and regasify the liquid air. In this example, superheated steam is injected into combustor 600 and mixed with pressurized fuel and gaseous air which burns to produce hot gas for expansion through turbine-generator 700. Exhaust gas from the turbine generator passes through heat recovery steam generator 800, which converts boiler feedwater from pump 930 into superheated steam. The exhaust gas then enters regasifier 400 where it evaporates and heats high pressure liquid air. Water vapor in the exhaust gas, which includes water of combustion and injected steam, is condensed by heat transfer in the regasifier, separated by condensate collector 920, and transferred to condensate tank 940 for subsequent treatment by water treatment system 950, after which it is transferred to treated water tank 960. Depending on the amount of injected steam and the amount and composition of the fuel there may be excess water recovered or there may be a need for makeup water.

By way of illustration, in one example about 15 kg/s of boiler feedwater enters heat recovery steam generator 800 at about 318 K, where it is evaporated at about 33 bar and superheated to about 853 K. The resulting superheated steam is mixed in combustor 600 with about 100 kg/s of re-gasified air, and burned completely with about 3.69 kg/s of methane. The hot combustion gas enters turbine-generator 700 at a temperature of about 1675 K and a pressure of about 31 bar, where it generates about 135.2 MW of electric power. The exhaust gas from turbine-generator 700 enters heat recovery steam generator 800 at a temperature of about 876 K and leaves at about 528 K, then enters regasifier 400 to evaporate the liquid air at about 35 bar and then heat the air to about 513 K. Flue gas leaves gasifier 400 at a temperature of about 327 K, and about 14.2 kg/s of condensate is collected by collector 920, so about 95% of the injected steam would be recovered. If air liquefaction consumes about 0.4 kWh per kilogram, then the Round Trip Efficiency (RTE, ratio of electric power produced to consumed) is about 94% and the Heat Rate (specific fuel consumption) is about 5460 kJ/kWh.

Increasing the steam injection flow rate increases the power produced by turbine-generator 700 and increases the Round Trip Efficiency, at the expense of increased Heat Rate and an increased requirement for makeup water. Assuming the same turbine inlet pressure (about 31 bar), temperature (about 1675 K) and steam injection temperature (about 853 K, which may be limited by metallurgical constraints), with about 20 kg/s of steam injection the fuel flow increases to about 4.1 kg/s and the power output increases to about 145.6 MW. With the same electric power requirement for air liquefaction, the Round Trip Efficiency becomes 100% at a Heat Rate of about 5656 kJ/kWh, while the water recovery ratio is reduced to about 78%.

Likewise, decreasing the steam injection flow rate, at the same turbine inlet and steam injection pressure and temperature, has an opposite effect. With about 10 kg/s of steam injection, the fuel flow rate is reduced to about 3.275 kg/s and the power output decreases to about 126.0 MW, resulting in a Round Trip Efficiency of about 87% and a Heat Rate of about 5220 kJ/kWh, while the water recovery ratio becomes about 128%, since almost all of the water of combustion is recovered as condensate.

FIG. 5 illustrates the effect on efficiency as the steam injection ratio is increased. The Round Trip Efficiency as defined above increases in proportion to steam injection because the additional mass flow increases the output power. The water recovery ratio decreases with increasing steam injection to a minimum that occurs when there is sufficient flow of feedwater to cause condensation in the heat recovery steam generator. The Fuel Efficiency (power output divided by the fuel heat release rate) and Total Efficiency (power divided by the fuel and electric power input) also decrease to a minimum and then increase with increasing steam injection. The maximum steam to air ratio occurs when all of the oxygen has been consumed, at about 50% steam to air ratio under the temperature and pressure conditions described above.

The above results illustrate the trade-off between Round Trip Efficiency and power output on the one hand versus Heat Rate and water recovery ratio on the other hand. Liquid air power and storage systems and plants as described herein may be designed accordingly, considering the costs of fuel and water, and the relative value of consumed electricity used for liquefying air compared to generated electricity. Once designed, a liquid air and storage system may be operated with different steam injection flow rates, depending on the contemporaneous economic factors. Operation at conditions different from the design condition may be readily accommodated by varying the inlet pressure to the turbine, including by adjustment of the discharge pressure of liquid air pump 300 and boiler feedwater pump 930.

Condensing water vapor from the exhaust gas can consume a significant fraction of the heat sink provided by the liquid air in a liquid air power and storage system. As an alternative, steam injection into the combustor may be reduced or eliminated to reserve an increased portion of the heat sink provided by the liquid air for freezing and separating carbon dioxide out of the combustion turbine exhaust gas.

For example, in the liquid air power and storage system illustrated in FIG. 6 pressurized liquid air from pump 300 is split into two streams. One stream of pressurized liquid air is gasified in regasifier 400 by heat exchange with and condensation of the exhaust from organic Rankine Cycle turbine generator 1000. Condensed organic working fluid exiting regasifier 400 is pressurized by pump 1030, evaporated and superheated in organic vapor generator 1010 by hot exhaust gas exiting turbine-generator 700, and expanded again through turbine-generator 1000 in an organic Rankine bottoming cycle operating similarly to those described above.

Water of combustion in the exhaust gas from turbine generator 700 is condensed by heat transfer in organic vapor generator 1010 and removed by collector 920. The exhaust gas then flows into carbon dioxide freezer 1100, which is cooled by heat transfer to the second stream of pressurized liquid air, causing formation of dry ice (frozen carbon dioxide) which is removed from the exhaust gas by carbon dioxide separator 1110. The frozen carbon dioxide may be collected on a scraped-surface heat exchanger or collected as snow in an inertial separator, for example. Though shown separately, carbon dioxide freezer 1100 and carbon dioxide separator 1110 may be an integrated unit. The collected frozen carbon dioxide may, for example, be sequestered or used as a heat sink to reduce energy consumption of the air liquefaction unit. From carbon dioxide freezer 1100 and carbon dioxide separator 1110, the exhaust gas flows to optional cold storage unit 1120 where the cold exhaust gas cools a storage medium prior to being discharged up the flue. Cold storage unit 1120 may provide cooling to air liquefaction unit 100 during the charging phase of operation. The storage medium in cold storage unit 1120 may comprise, for example, chilled water, brine, a silicate ester (e.g., Coolanol®), a water-ice mixture, or any other suitable material.

The second stream of pressurized liquid air may be partially or completely regasified in carbon dioxide freezer 1100, from which it flows to regasifier 400 where it is mixed with the first stream of liquid air to assist in cooling the organic working fluid exhaust from turbine-generator 1000. The combined air stream then flows through optional auxiliary air heater 410 before being mixed and combusted with pressurized fuel in combustor 600. The resulting combustion gases, including the aforementioned water of combustion and carbon dioxide, expand through turbine-generator 700 to produce power.

By way of illustration, about 100 kg/s of liquid air at a pressure of about 32 bar and a temperature about 499 K mixes with about 3.06 kg/s of methane and burns in combustor 600 to produce a hot gas mixture at about 1675 K and about 31 bar. This mixture expands through turbine-generator 700 to produce about 108 MW of electric power, exhausting at about 848 K. The exhaust gas enters organic vapor generator 1010 where it is cooled by the organic working fluid to about 273 K, transferring about 83 MW of heat.

About 6.8 kg/s of the condensed water of combustion is removed from the exhaust gas by collector 920. The remaining 96.2 kg/s of exhaust gas flows to carbon dioxide freezer 1100, which transfers about 7.7 MW of heat to liquid air, and which cools the exhaust gas to about 193 K. About 8.4 kg/s of dry ice (solid carbon dioxide) is removed by carbon dioxide separator 1110, and the remaining 87.8 kg/s of exhaust gas flows through cold storage unit 1120, absorbing about 8.9 MW of heat to warm the flue gas to about 293 K.

About 20 percent of the liquid air is directed to carbon dioxide freezer 1100. About 53.5 MW of heat is absorbed in regasifier 400 to condense about 50 kg/s of propane at about 0.125 bar and then cool the propane to about 173 K. The liquid propane is pumped to about 105 bar and then flows to the organic vapor generator 1010 where it is evaporated and then superheated to a temperature of about 726 K. The superheated propane vapor flows through the organic Rankine cycle turbine-generator 1000 to generate about 31 MW of power. The low pressure propane at about 498 K then flows to regasifier 400 to be cooled by about 20 kg/s of cold air from the carbon dioxide freezer 1100 and the liquid air.

Under these conditions, a total of about 136.8 MW of electric power is generated by the two generators, with a fuel consumption of about 4648 kJ/kWh. The initial charge of liquid air may consume for example about 0.4 kWh of electricity per kg of liquid air, resulting in a Round Trip Efficiency (as defined to refer only to electrical power) of about 96% and an overall efficiency (including fuel consumption) of about 43.6%. Use of the optional cold storage unit during subsequent charging cycles would reduce the electric power consumption and increase the efficiency.

If high carbon fuels such as coal, petroleum coke, or biomass were burned instead of methane, more liquid air would be directed to the carbon dioxide freezer to effect the separation of carbon dioxide.

The availability and cost of electricity may vary hourly, daily and seasonally, as may the load demand and price of electricity. These parameters are influenced by market and regulatory forces, as well as by constraints on the operation of generation and transmission assets due to weather, emissions, planned or forced outages, and public policy. Accordingly, the quantity of air to be liquefied and stored may be varied throughout each day and seasonally.

FIG. 7 shows a plot that illustrates a diurnal charge/discharge cycle for an energy storage system with a 50% Round Trip Efficiency when the charging and discharging are not coincident. In the example illustrated, 40 MW of electric power charges the storage for 20 hours. The storage is discharged over 4 hours at a rate of 100 Megawatts. The RTE is the ratio of energy discharge to energy charge, or 400 MW-hour over 800 MW-hour. For non-coincident charging and discharging, Π=η/γ, where H is the ratio of discharge power Pd to charge power Pc in MW, η is the ratio of discharge energy to charge energy in MW-hours, and γ is the ratio of discharge time to charge time in hours.

It may be advantageous for the charging system to operate continuously, as illustrated by the plot shown in FIG. 8. In order to achieve the same net power Pn during discharge as in the non-coincident example, the discharge power must increase compared to the non-coincident example. It can be shown that Pn=Pd [1−γ/η)] (where Pd, γ, and η are defined above). In the example illustrated in FIG. 4, 50 MW of power is used to charge the system 24 hours per day. To deliver 100 MW of net power during the 4-hour long discharge period, the discharge rate is 150 MW. The overall energy consumption increases from 800 to 1200 MW-hours.

LAPS systems as described herein are typically more efficient than conventional combustion turbine-generators (CTGs). In addition, LAPS systems may be operated at reduced load (e.g., at less than the turbine generator's nameplate capacity) without significantly reducing the efficiency of power production. This is advantageous compared to conventional combustion turbine generators operating in an open Brayton Cycle.

Referring now to FIG. 9, a CTG 1200 configured to operate in an open Brayton Cycle comprises a compressor section 1210, a combustion section 1220, and a power (turbine) section 1230. The power section drives the compressor via a common shaft, with the CTG net output power being the power produced by the turbine less the power consumed by the compressor. Cooling air is extracted from the compressor to cool the hot parts in the turbine section. For electric power generation applications, the shaft is also coupled to a generator (not shown), which rotates at a fixed frequency, set by the requirements of the electrical grid. For example 3600 rpm or 60 Hertz in North America, or a multiple based on the number of generator poles. The CTG shaft speed will still be set by the electrical grid, even when a gearbox is interposed between the CTG shaft and the generator.

At fixed speed, the compressor section of the CTG sweeps a constant volumetric flow rate of air, whatever the power delivered to the generator. For a given air temperature entering the turbine, the mass flow is approximately fixed, and except for cooling air is constant throughout the machine. Both the compressor and turbine follow adiabatic processes, so work done by the compressor and turbine is given by:

$W = {\frac{kRT}{k - 1}\left\lbrack {{\Pi\;\frac{k - 1}{k}} - 1} \right\rbrack}$ where

W is the Work

k is the specific heat ratio

T is the temperature at the beginning of the process

R is the gas constant

Π is the pressure ratio

Assuming the pressure drop across the combustion section is small, the pressure ratio is approximately the same for both adiabatic processes. The turbine produces more work than required by the compressor because the turbine's inlet air temperature is much higher than that of the compressor by virtue of the combustion heat addition.

For part load operation (e.g., operation at less than nameplate capacity) of a conventional CTG, the power output is reduced by reducing the fuel flow, thereby reducing the temperature of combustion gases entering the turbine. The fuel reduction also reduces the pressure ratio by changing the density of the combustion gases. This is because the turbine section is ‘choked’ for pressure ratios more than about 2:1, meaning that the pressure drop is proportional to the volumetric flow rate through the turbine. But reducing the turbine inlet temperature for part load operation also reduces the efficiency of the CTG.

FIG. 10 shows the effect of part load operation for critical parameters for a conventional single-shaft gas turbine compared to the design point (nameplate capacity) full-load at ISO conditions. Part load conditions were estimated by reducing fuel flow and adjusting the pressure ratio to keep the volumetric flow rate through the turbine section unchanged (the choking condition). Component efficiency was assumed to be unchanged across the operating range, and the water injection flow rate was assumed to be proportionate to fuel flow. As can be seen in FIG. 10, the mass flow is essentially constant since the fuel and water injection flow rates are a small fraction of the air flow. At part load, the pressure and temperature at the turbine inlet each are reduced. The part load efficiency falls rapidly below about 60% of the full load rating. The Turbine Outlet Temperature (TOT) ratio also decreases as power output is reduced.

FIG. 11 shows an example simple LAPS configuration, similar to those described above, using turbine exhaust heat to regasify liquid air but without steam injection or bottoming cycles. The cryopump 300 increases the pressure of liquid air which flows through a gasifier 400, where the liquid evaporates. The gaseous air is mixed with fuel and steam in a burner 600, and then the hot pressurized air expands across the turbine-generator 700 to produce power. Pressurized cooling air is optionally supplied to the hot turbine section. Exhaust heat may be optionally transferred by the boiler to produce steam for injection (not shown). The water of combustion is condensed and separated.

For a conventional CTG, typically more than 60% of the shaft work produced by the turbine section is consumed for air compression. So a conventional CTG rated for 47 MW net generator power output has a turbine that actually produces about 116 MW, with the compressor section consuming about 69 MW. Since LAPS systems as described herein pump a liquid, which is roughly incompressible, virtually no power is consumed to pressurize liquid air. Accordingly, in the LAPS systems all of the power section shaft work can be delivered to the generator. For the same power section, with identical flow rates and pressure ratios, the gross power at the generator terminals is much larger: for example about 114 MW for LAPS compared to about 47 MW for the conventional CTG. Consequently, LAPS systems have much higher fuel efficiency than conventional CTGs. For example, a conventional CTG as shown in FIG. 9 might have a fuel efficiency of 42% at full load, whereas the same power section in the LAPS steam injection configuration may have for example a fuel efficiency of 63%.

Refrigeration work is required to liquefy the air, but the LAPS cycle de-synchronizes the work of delivering high pressure air to the turbine from the extraction work to energy storage by the liquefaction process. Equally important is that LAPS decouples the source and cost of the power used to produce the stored energy from the fuel cost of generating power, thereby providing an arbitrage mechanism to shift plentiful and low cost power to periods of higher demand and higher value.

As noted above, the LAPS systems described herein enable improved (more efficient) part load operation compared to conventional CTGs, or Liquid Air Energy Storage Systems operated like conventional CTGs. This is because in operation of a LAPS system the liquid air flow rate may be adjusted proportional to load and the fuel flow rate may be adjusted to maintain the maximum allowable temperature at the inlet and outlet of the turbine. The liquid air flow can be varied in response to a load signal by means of a variable speed drive on the cryopump or by means of a control valve, for example. Fuel flow can be varied in response to a signal based on the Turbine Outlet Temperature, for example. In contrast and as described above, in part load operation of a conventional single-shaft CTG the volumetric and mass air flow rates through the turbine remain approximately constant and the turbine inlet temperature and therefore the efficiency drops rapidly with decreasing load.

FIG. 12 shows performance of the example simple LAPS configuration of FIG. 11 at a range of loads. Because less work is extracted as the turbine pressure ratio is reduced, the Turbine Outlet Temperature (TOT) becomes the limiting factor, and necessitates a reduction in fuel flow, and hence a reduction of the Turbine Inlet Temperature (TIT). If the materials at the outlet of the turbine were similarly heat resistant or cooled as those at the inlet, this constraint could be relaxed, and efficiency would be almost constant throughout the load range.

The design point power output of the turbine section in a conventional CTG is typically constrained by the turbine inlet temperature limit and/or the compressor surge limit. Because the LAPS system does not have a compressor surge limit, it is practical to overload the turbine section and boost power output by increasing the turbine mass flow and pressure ratio, while still observing the inlet temperature limit. The turbine mass flow may be increased beyond the nominal design case by, for example, increasing air flow and/or adding steam.

FIG. 13 shows another example LAPS configuration, similar to that of FIG. 11, using steam injection to the burner and optionally to the turbine. Condensate, which may include both the injected steam and water of combustion, is separated from the turbine exhaust gas stream and recycled to a boiler 800 through a water treatment system 950. The water treatment system may condition the condensate and add makeup water, and may include intermediate water storage tanks. Steam produced in the boiler is injected into the burner. Optionally, additional steam produced in the boiler is used to cool hot turbine components. Depending on the quantity of steam injection, excess condensate may be produced. Water flow into the boiler can be varied in response to the steam temperature exiting the boiler, for example, optionally with a bias or feedforward signal from the air flow or demand signal.

FIG. 14 shows that performance of a LAPS configuration using steam injection for power augmentation or NOx control may be similar to performance without steam injection.

Among the advantages that the LAPS systems described herein may provide, in addition to maintaining the highest efficiency across the broad load range, is the efficiency of using liquid air. From an operational perspective, a LAPS system could operate economically at low loads for extended periods of time to provide regulation and voltage support services to the grid, also known as ancillary services, in addition to energy storage, capacity and load following.

The marginal fuel cost is proportional to Heat Rate. The economic advantages of the LAPS systems described herein can be inferred from FIG. 15, which compares the Heat Rate versus power output of two example LAPS systems (one utilizing steam injection) and an example conventional CTG system with the same turbine section. For the conventional CTG system, three 47 MW nominal rating CTGs are assumed to share load, with units shutdown in sequence as load decreases first below 94MW and then below 47MW. The plots in FIG. 15 demonstrate that the LAPS Heat Rate is always substantially less than the CTG Heat Rate throughout the load range.

As described above, a LAPS system may include a bottoming cycle. Because the turbine outlet temperature of the upper cycle (using the regasified liquid air) in such a combined cycle LAPS system may be approximately constant with variations in load, the heat transfer conditions to the bottoming cycle may be maintained approximately constant by varying the flow rate in the bottoming cycle roughly in proportion to the flow rate of liquid air. A bottoming Rankine Cycle may therefore operate in sliding pressure mode, maintaining the throttle temperature at the inlet to a bottoming cycle turbine to optimize the part load efficiency of the bottoming cycle. This will also reduce thermal stresses that would otherwise result from reduced Turbine Outlet Temperature during part load operation as in a conventional CTG.

When a synchronous alternating current machine is at zero torque (i.e, neither driven as a generator, nor driving a load as a motor), and is over-excited, it is termed a synchronous condenser. Synchronous condensers provide reactive power measured with a unit of VARs (Volt-Ampere Reactive) in contrast to what is termed ‘real’ power, measured in Watts (which could be termed Volt-Ampere Resistive). Reactive power is required for voltage support on the grid, and a synchronous condenser is one means of adjusting and controlling voltage. The excitation of loaded generators can be adjusted, but sometimes it is necessary to adjust reactive power independently of the real power, especially if the generator is far from the load. A combustion turbine peaking plant could provide this service, but because there is substantial fuel consumption to spin the compressor in addition to the generator, this would be an exceptional use. On the other hand the LAPS machine does not drive a compressor, so the torque requirements at idle condition are very low, with only bearing friction and windage losses to overcome. Accordingly, it is feasible to consider operating a LAPS machine to provide reactive power, essentially a synchronous condenser, with very low or zero power output. Because very little air and fuel is needed at the idle condition to rotate the turbine and generator, the LAPS machine could operate for extended periods as a synchronous condenser. It is also feasible to motor the turbine and generator, that is consume a small amount of electricity to spin the system, as in a conventional synchronous condenser, without consuming any air or fuel.

During part load operation cooling air requirements for the hot section of the LAPS system can be optimized by increasing the cooling air flow, even as the air flow leaving the combustor is reduced. This may be done by using a control valve to manage the flow of cooling air that bypasses the combustor. This is in contrast to a conventional CTG, where cooling air flow would decrease at part load because of lower compressor discharge pressure.

FIG. 16 shows that the cooling air can be drawn from an intermediate location in the gasifier before the regasified air has reached maximum temperature, in order to provide more effective cooling. This may also improve the efficiency of the LAPS system, because the remaining air will receive all of the heat and therefore enter the combustor at higher temperature, which accordingly reduces the fuel needed to heat the combustion gases to the specified turbine inlet temperature. Alternatively, the cooling air could be regasified in a separate circuit.

As described earlier cooling could be provided by saturated or slightly superheated steam instead of or in combination with cooling air. The steam has a higher heat capacity than air, so its cooling effectiveness is correspondingly higher.

Liquid air energy storage systems as described herein may provide numerous technical, economic, and practical benefits and advantages over the prior art. These benefits may include the following.

-   -   Hybridization with a fuel source leverages the work in the         stored liquid air, permitting a smaller liquefaction plant and         smaller tank farm for cryogenic storage.     -   Flexible combustion system configurations permit a wide variety         of fuels to be burned, and permit greatly reduced nitrogen oxide         pollutants by means of staging the combustion with cooling         superheated steam and diluent air.     -   The use of a smaller turbine, adapted from existing aero         derivative or industrial frame units, because all of the shaft         work produces useful power. In contrast a conventional         combustion turbine generator delivers only about one-third of         the shaft power to electricity generation, with most shaft work         used for air compression.     -   The power output is not significantly influenced by atmospheric         pressure or temperature, unlike conventional turbine generators         which suffer significant performance degradation at high         altitudes or high ambient temperatures.     -   The Heat Rate of the unit (KJ of fuel per KW-HR of electricity)         may be superior to other fossil thermal generation systems, and         consume for example less than half the fuel of a gas turbine         peaker plant.     -   The use of a compact once through heat recovery steam generator         with regenerative steam injection may permit faster startup,         more flexible operation, and better part load efficiency than         conventional combined cycle power plants.     -   Water may be condensed from the flue gas for recycling in the         Heat Recovery Steam Generator and excess water can be produced,         potentially providing additional benefit.     -   Carbon dioxide can be frozen from the flue gas, making the         system carbon neutral, or better considering the carbon dioxide         that may be removed during liquefaction.

The marginal generating cost of electricity with the liquid air energy storage systems described herein is the cost of fuel and electricity. Using the process conditions of Tables 1-5 and assuming four hours of storage, an example system would generate about 463 Megawatt-hours and consume about 584 Megawatt-hours of electricity and about 672.8 Megawatt-hours of fuel (Higher Heating Value). A peaker plant, such as a GE LM6000 PC-SPRINT at full-load with a heat rate of about 9163 Btu/kw-hr (Higher Heating Value) would consume about 4242 Million Btu to produce 463 Megawatt-hours of electricity (Gross, before accounting for auxiliary load). For fuel at about $6.00 per Million Btu, the peaker plant fuel cost is about $25,455 compared to about $12,000 for the fuel consumed by the invention. For a marginal cost, the invention could spend $13,455 to purchase 584 Megawatt-hours of electricity, for liquefaction costs. This is about $23 per megawatt-hour. Consideration should also be given to the cost of carbon dioxide emissions. Burning natural gas, the LM6000 SPRINT PC produces 0.5 tonnes per Megawatt-hour, compared to 0.26 tones per Megawatt-hour for the conditions described in tables 1-5. At a CO2 value of $30 per tonne, this is equivalent to $7.20 per MWH, which would make the break-even electricity cost for charging the storage system to be about $30 per Megawatt-Hour. Off-peak electricity is often less expensive than this.

The following enumerated clauses provide additional non-limiting aspects of the disclosure.

1. A method of operating a liquid air power and storage system comprising a combustor, a turbine, and a generator, the method comprising:

regasifying liquid air or liquid air components to produce gaseous air or gaseous air components;

combusting at least a portion of the gaseous air or gaseous air components in a combustor with a gaseous fuel to form a gaseous primary working fluid at an elevated temperature;

expanding the primary working fluid through the turbine;

driving the generator with the turbine to generate electricity; and

adjusting a mass flow rate of the gaseous air or gaseous air components into the combustor proportionally to a changing load on the turbine from the generator.

2. The method of clause 1, wherein regasifying the liquid air or liquid air components to produce gaseous air or gaseous air components comprises regasifying the liquid air or liquid air components to produce gaseous air or gaseous air components using heat from an exhaust gas stream from the turbine;

3. The method of clause 1 or clause 2, comprising adjusting a mass flow rate of the fuel into the combustor to maintain a desired temperature of the primary working fluid at an inlet to the turbine.

4. The method of clause 1 or clause 2, comprising adjusting a mass flow rate of the fuel into the combustor to maintain a desired temperature of the primary working fluid at an outlet to the turbine.

5. The method of any of clauses 1-4, comprising operating the generator as a synchronous condenser.

6. The method of any of clauses 1-5, wherein the mass flow rate of the gaseous air or gaseous air components into the combustor is less than that required to operate the turbine at nameplate capacity, a mass flow rate of the fuel into the combustor is less than that required to operate the turbine at nameplate capacity, and a temperature of the primary working fluid at an inlet to the turbine measured in degrees Kelvin is at least 60% of that required to operate the turbine at nameplate capacity.

7. The method of clause 6, wherein the temperature of the primary working fluid at the inlet to the turbine measured in degrees Kelvin is at least 65%, at least 70%, at least 75%, at least 80%, at least 85%, at least 90%, at least 95%, or at least 100%.

8. The method of clause 6 or clause 7, wherein the mass flow rate of the gaseous air or gaseous air components into the combustor is less than 95%, less than 90%, less than 85%, less than 80%, less than 75%, less than 70%, less than 65%, less than 60%, less than 55%, less than 50%, less than 45%, less than 40%, less than 35%, less than 30%, less than 25%, less than 20%, less than 15%, less than 10%, or less than 5% of that required to operate the turbine at nameplate capacity.

9. The method of any of clauses 1-4, wherein the mass flow rate of the gaseous air or gaseous air components into the combustor is greater than that required to operate the turbine at nameplate capacity.

10. The method of any of clauses 1-9, comprising heating water to produce steam using heat from an exhaust gas stream from the turbine, and mixing some or all of the steam with the gaseous air or gaseous air components and gaseous fuel during combustion of the gaseous fuel.

11. The method of any of clauses 1-10, comprising heating water to produce steam using heat from an exhaust gas stream from the turbine, and using some or all of the steam to cool the turbine.

12. The method of any of clauses 1-11, comprising cooling the turbine with a portion of the gaseous air or gaseous air components provided to the turbine separately from the primary working fluid.

13. The method of clause 12, wherein the portion of the gaseous air or gaseous air components provided to the turbine is drawn from an intermediate location in a gasifier at a temperature lower than a maximum temperature of the gaseous air or gaseous air components at an outlet of the gasifier.

14. The method of clause 12 or clause 13, comprising decreasing the mass flow rate of the gaseous air or gaseous air components into the combustor in combination with increasing a flow rate of the portion of gaseous air or gaseous air components provided to the turbine separately from the primary working fluid to cool the turbine.

15. The method of any of clauses 1-14, comprising heating a secondary working fluid with heat from an exhaust gas stream from the turbine to convert the secondary working fluid from a liquid phase to a gas phase, and using the heated working fluid in a bottoming Rankine power cycle to generate additional electricity.

16. The method of clause 15, comprising proportionately varying a flow rate of the secondary working fluid and the mass flow rate of the gaseous air or gaseous air components into the combustor.

17. The method of clause 15, comprising operating the bottoming Rankine power cycle in sliding pressure mode.

18. The method of any of clauses 15-17 wherein the secondary working fluid is an organic working fluid, comprising regasifying the liquid air or liquid air components using heat transferred from the organic secondary working fluid to the liquid air or liquid air components, thereby condensing the organic secondary working fluid from a gas phase to a liquid phase.

19. The method of any of clauses 15-17, wherein the secondary working fluid is or comprises water.

20. The method of any of clauses 1-19, comprising producing the liquid air or liquid air components in an electrically powered liquefaction process and storing the liquid air or liquid air components for later regasification and use in combusting the gaseous fuel.

21. The method of any of clauses 1-20, comprising catalytically combusting the gaseous air or gaseous air components with the gaseous fuel.

22. The method of any of clauses 1-21, wherein the gaseous fuel is or comprises natural gas, hydrogen, or natural gas and hydrogen.

23. The method of any of clauses 1-22, comprising separately providing different ones of the gaseous air components to a combustor in which they are combusted with the fuel.

24. The method of any of clauses 1-23, comprising providing the gaseous primary working fluid to an inlet of the turbine at a temperature greater than or equal to about 1600 K and a pressure greater than or equal to about 30 bar.

25. The method of any of clauses 1-24, comprising freezing carbon dioxide out of an exhaust gas stream from the turbine by transferring heat from the turbine exhaust gas stream to the liquid air or liquid air components to cool the turbine exhaust gas stream.

26. The method of any of clauses 1-25, comprising condensing water out of an exhaust gas stream from the turbine thereby increasing the turbine power output.

This disclosure is illustrative and not limiting. Further modifications will be apparent to one skilled in the art in light of this disclosure and are intended to fall within the scope of the appended claims.

TABLE 1 Air flow to and from storage Stream Inlet air Liq Air from storage HP Liq Air Combustor Inlet Air Burner Inlet Air Unit Pressure 1 1 35 34.975 34.974 bar Temperature 293 78.6574 80.5105 375.223 375.223 K Flow rate 101.36 101.36 101.36 101.36 101.36 kg/s Flow Oxygen 23.4564 23.4564 23.4564 23.4564 23.4564 kg/s Flow Nitrogen 76.5471 76.5471 76.5471 76.5471 76.5471 kg/s Flow Water 0 0 0 0 0 kg/s Flow Argon 1.30567 1.30567 1.30567 1.30567 1.30567 kg/s Flow Carbon dioxide 0.0483579 0.0483579 0.0483579 0.0483579 0.0483579 kg/s Flow Neon 0.00128381 0.00128381 0.00128381 0.00128381 0.00128381 kg/s Flow Helium-4 7.3402e−05 7.3402e−05 7.3402e−05 7.3402e−05 7.3402e−05 kg/s Flow Krypton 0.000334303 0.000334303 0.000334303 0.000334303 0.000334303 kg/s Flow Xenon 0.000229717 0.000229717 0.000229717 0.000229717 0.000229717 kg/s Flow Nitrous oxide 0.000452814 0.000452814 0.000452814 0.000452814 0.000452814 kg/s Flow Methane 0.000112281 0.000112281 0.000112281 0.000112281 0.000112281 kg/s Vapor phase Mass fraction 1 0.000835267 1 1 Liquid phase Mass fraction 0.999165 1

TABLE 2 Combustor streams (assumes no nitrogen oxide formation in combustor 600) Stream Fuel, Air, Steam Organic Vapor Mixture entering CTG (700) HRSG (800) Generator Combustor (600) Hot Gas Exhaust Exhaust (1010) Exhaust Flue Gas Unit Pressure 31 31 1.5 1.25 1 1 bar Temperature 458.865 1473.16 809.466 571.78 310.828 310.828 K Flow rate 113.965 113.965 113.965 113.965 113.965 101.198 kg/s Flow Oxygen 23.4564 11.3689 11.3689 11.3689 11.3689 11.3688 kg/s Flow Nitrogen 76.5471 76.5471 76.5471 76.5471 76.5471 76.5471 kg/s Flow Water 9.57533 16.3806 16.3806 16.3806 16.3806 3.61352 kg/s Flow Argon 1.30567 1.30567 1.30567 1.30567 1.30567 1.30567 kg/s Flow Carbon 0.0484588 8.36073 8.36073 8.36073 8.36073 8.36059 kg/s dioxide Flow Neon 0.00128381 0.00128381 0.00128381 0.00128381 0.00128381 0.00128381 kg/s Flow Helium-4 7.34021e−05 7.34021e−05 7.34021e−05 7.34021e−05 7.34021e−05 7.3402e−05 kg/s Flow Krypton 0.000334303 0.000334303 0.000334303 0.000334303 0.000334303 0.000334303 kg/s Flow Xenon 0.000229717 0.000229717 0.000229717 0.000229717 0.000229717 0.000229717 kg/s Flow Nitrous 0.00045282 0.00045282 0.00045282 0.00045282 0.00045282 0.000452812 kg/s oxide Flow Methane 3.03011 0 0 0 0 0 kg/s Vapor phase 1 1 1 1 0.887973 1 Mass fraction Liquid phase 0.112027 0 Mass fraction

TABLE 3 Water streams Stream Excess (Produced) Water Condensed Water Boiler Feedwater Superheated Steam Unit Pressure 1 1 33 31 bar Temperature 310.828 310.828 311.114 809.467 K Flow rate 3.1918 9.57541 9.57541 9.57544 kg/s Flow Oxygen 1.96165e−06 5.88496e−06 5.88496e−06 5.88496e−06 kg/s Flow Nitrogen 7.01679e−07 2.10504e−06 2.10504e−06 2.10503e−06 kg/s Flow Water 3.19177 9.5753 9.5753 9.57533 kg/s Flow Argon 2.15592e−07 6.46776e−07 6.46776e−07 6.46775e−07 kg/s Flow Carbon dioxide 3.36205e−05 0.000100862 0.000100862 0.000100862 kg/s Flow Neon 1.01417e−10 3.04252e−10 3.04252e−10 3.04252e−10 kg/s Flow Helium-4 6.36421e−12 1.90926e−11 1.90926e−11 1.90926e−11 kg/s Flow Krypton 1.04047e−10 3.12141e−10 3.12141e−10 3.12141e−10 kg/s Flow Xenon 1.37266e−10 4.11798e−10 4.11798e−10 4.11797e−10 kg/s Flow Nitrous oxide 1.91954e−09 5.75863e−09 5.75863e−09 5.75865e−09 kg/s Flow Methane 0 0 0 0 kg/s Vapor phase Mass fraction 1 Liquid phase Mass fraction 1 1 1

TABLE 4 Organic fluid flows (shown as propane) Stream Low Pressure High Pressure High Pressure Low Pressure Organic Liquid Organic Liquid Organic Vapor Organic Vapor Unit Pressure 1.5 101.5 101.25 1.5 bar Temperature 200 205.619 571.78 419.844 K Flow rate 60 60 60 60 kg/s Vapor phase Mass 1 1 fraction Liquid phase Mass 1 1 fraction

TABLE 5 Principal power flows Operation Value Unit Liquid Air Storage (200) 145.958 MW Combustor (600) 168.2 MW Auxiliary Air Heater (410) 0.0 MW Startup Heater (1020) 0.0 MW Auxiliary Loads (pumps) 4.4839 MW Combustion Turbine Generator (700) 102.302 MW Organic Turbine-Generator (1010) 18.0428 MW Total Efficiency (HHV) 38.31 % Round Trip Efficiency (electric) 82.45 % 

What is claimed is:
 1. A method of operating a liquid air power and storage system comprising a combustor, a turbine, and a generator, the method comprising: regasifying liquid air or liquid air components to produce gaseous air or gaseous air components using heat from an exhaust gas stream from the turbine; combusting at least a portion of the gaseous air or gaseous air components in the combustor with a gaseous fuel to form a gaseous primary working fluid at an elevated temperature; expanding the primary working fluid through the turbine; driving the generator with the turbine to generate electricity; adjusting a mass flow rate of the liquid air or liquid air components from a first non-zero rate to a second non-zero rate proportionally to a changing non-zero mechanical load on the turbine from the generator wherein the liquid air or the liquid air components are regasified to correspondingly adjust a mass flow rate of the gaseous air or gaseous air components into the combustor from a first non-zero rate to a second non-zero rate; and adjusting a mass flow rate of fuel into the combustor to compensate for the adjusting of the mass flow rate of the liquid air or liquid air components from the first non-zero rate to the second non-zero rate proportionally to the changing non-zero mechanical load on the turbine from the generator and thereby maintain a temperature at an air inlet to the turbine or a temperature at an exhaust outlet of the turbine at or below a limit temperature wherein the first non-zero rate and the second non-zero rate of the gaseous air or gaseous air components into the combustor and the mass flow rate of the fuel into the combustor are less than that required to operate the turbine at nameplate capacity.
 2. The method of claim 1, comprising operating the generator as a synchronous condenser.
 3. The method of claim 1, wherein a temperature of the primary working fluid at an inlet to the turbine measured in degrees Kelvin is at least 60% of that required to operate the turbine at the nameplate capacity.
 4. The method of claim 1, comprising heating water to produce steam using heat from an exhaust gas stream from the turbine, and mixing some or all of the steam with the gaseous air or gaseous air components and gaseous fuel during combustion of the gaseous fuel.
 5. The method of claim 1, comprising heating water to produce steam using heat from an exhaust gas stream from the turbine, and using some or all of the steam to cool the turbine.
 6. The method of claim 1, comprising cooling the turbine with a portion of the gaseous air or gaseous air components provided to the turbine separately from the primary working fluid.
 7. The method of claim 6, wherein the portion of the gaseous air or gaseous air components provided to the turbine is drawn from an intermediate location in a gasifier at a temperature lower than a maximum temperature of the gaseous air or gaseous air components at an outlet of the gasifier.
 8. The method of claim 6, comprising decreasing the mass flow rate of the gaseous air or gaseous air components into the combustor in combination with increasing a flow rate of the portion of gaseous air or gaseous air components provided to the turbine separately from the primary working fluid to cool the turbine.
 9. The method of claim 1, comprising heating a secondary working fluid with heat from an exhaust gas stream from the turbine to convert the secondary working fluid from a liquid phase to a gas phase, and using the heated working fluid in a bottoming Rankine power cycle to generate additional electricity.
 10. The method of claim 9, comprising proportionately varying a flow rate of the secondary working fluid and the mass flow rate of the gaseous air or gaseous air components into the combustor.
 11. The method of claim 9, comprising operating the bottoming Rankine power cycle in sliding pressure mode.
 12. The method of claim 9, wherein the secondary working fluid is an organic working fluid, comprising regasifying the liquid air or liquid air components using heat transferred from the organic secondary working fluid to the liquid air or liquid air components, thereby condensing the organic secondary working fluid from a gas phase to a liquid phase.
 13. The method of claim 9, wherein the secondary working fluid is or comprises water.
 14. The method of claim 1, comprising catalytically combusting the gaseous air or gaseous air components with the gaseous fuel.
 15. The method of claim 1, comprising providing the gaseous primary working fluid to an inlet of the turbine at a temperature greater than or equal to about 1600 K and a pressure greater than or equal to about 30 bar.
 16. The method of claim 1, comprising freezing carbon dioxide out of an exhaust gas stream from the turbine by transferring heat from the turbine exhaust gas stream to the liquid air or liquid air components to cool the turbine exhaust gas stream. 